WO 2014/007727 A1 [1], a publication dated Jan. 9, 2014, of international patent request PCT/SE2013/050780 makes known the ‘Actuator for axial displacement of a gas-exchange valves of internal combustion engines’. According to [1], the actuator comprises a casing attached to the cylinder head of the engine, with a hollow internal cylinder open at the bottom end and containing an axially aligned two-way moving piston whose upper part connects to a piston rod axially aligned with the open hollow cylinder. The piston with its piston rod comprise the driven part of the actuator, the piston rod being made of two parts—an upper, thicker part fitting tightly to an opening in the actuator casing, and a lower, thinner part connecting the thicker part to the piston in the cylinder. The head part of the thick portion of the piston rod contacts a chamfered portion of the hollow recess space co-axial to the hollow cylinder. A diametrical opening is formed in this part of the piston rod, and a spring-loaded return valve located in an internal chamber in the piston closes off an axial opening along the rod axis. The inner chamber of the piston contains also a positioning piston reciprocating along its axis relative to the actuator piston, and contains the check-valve spring. The positioning piston interacts with the internal combustion engine valve stem which, moving through the internal combustion engine, enters the actuator cylinder. A conventional spring returns the opened internal combustion engine valve to its closed position.
The actuator case includes a circuit for controlled filling of pressurised gaseous fluid into the upper above-piston volume, on the piston rod side, and for removal of the fluid from that upper part in order to act upon the actuator piston and ensure its movement. The fluid circuit is connected to a pressurised gaseous fluid source and to a discharge-fluid receiver. Direct or indirect controlled fluid filling and discharging is provided for, one embodiment option including an indirect electrically controlled valve in the actuator-control fluid circuit, and another option including directly controlled electrical magnet acting upon the actuator body. The actuator includes also additional primary and secondary hydraulic circuits. The primary additional circuit represents an inlet through a fluid (hydraulic oil) filled check valve connecting the space in the casing above the upper part of the piston rod with the space in the casing above the head part of the piston rod, occupied by the upper surface of the free end of the piston rod at its top position, and an outlet via the controlled valve of the fluid receiver. The second auxiliary circuit comprises the envisioned diametral and co-axial openings in the piston rod connected via the check valve to the internal chamber in the piston, whereby the fluid (the hydraulic oil) bypasses the positioning piston to reach the lower part of the cylinder volume below the actuator piston. Thus, a hydraulic braking device is formed in the casing above the bevelled end of the piston rod and in the space formed above it, reducing the actuator piston speed before the free end of the piston-rod contacts the upper surface of the space, acting as a mechanical limiting stop in the casing.
The main emphasis regarding the actuator described in [1] is placed upon the hydraulic braking device described above and intended to reduce the valve speed exactly before valve head contacts the valve seat in the engine cylinder, ensuring controlled closing motion intended to preserve the details, and to reduce wear and disharmonious operation. When the engine valve closes, the actuator rod contacts the mechanical stop within the actuator casing and, the valve head contacting its seat in order to ensure correct deceleration of the actuator piston rod when the valve closes. The mechanical contact with the stop in the casing determines the final stage of engine-valve deceleration, but the linear elongation of the valve induced by higher temperatures during operation means that the upper end of the piston rod will come into contact with the actuator casing space of the hydraulic braking device, and that the elongated engine valve will not close completely, which is inadmissible. This problem has been resolved by way of the envisaged positioning piston acting as a hydraulic compensator within the inner chamber of the lower portion of the actuator piston and contacting the engine valve.
The deceleration effect of the envisaged hydraulic braking device in the mechanism described in [1] is variable, which is explained as follows: Upon entry into the hydraulic braking device space formed in the casing above it, the upper part of the piston rod displaces the hydraulic fluid contained in this space. Sufficient clearance is required between the casing and the piston-rod end, to allow the fluid to leave the space, but the clearance should not be too large because in that case the decelerating effect will be lost. The decelerating effect begins shortly before the engine valve reaches its fully-closed position under the force of its spring. Since internal combustion engines operate within a wide frequency range of rotation, the resistance of the hydraulic decelerating device will change as the engine RPM changes, i.e. increasing RPM increases also the force required to overcome this resistance. The engine valve spring creates a constant valve-closing force. Since increasing the engine RPMs increases the force required to overcome the resistance of the hydraulic braking device, this changing force will have an opposite effect on the permanent force of the valve spring, i.e. the decelerating effect of the braking device will vary, causing longer valve-closing delay. The decelerating effect of the braking device will be increased by the return motion of the piston in the cylinder pushing the gas out from the actuator cylinder. As this pushing is caused also by the valve spring, via a constant section opening, a resistance will be created requiring higher overcoming force at higher RPMs. The piston resistance when emptying the actuator cylinder volume will counteract the closing of the engine valve along its entire length of travel toward the closed position, and since as soon as the hydraulic braking device is actuated, both forces will be totalized, i.e. the decelerating effect of the braking device will be augmented by the same effect caused by the moving actuator piston. The decelerating effect of the hydraulic braking device will also arise when the engine valve opens. Since the end of the piston rod has reached the mechanical stop in the space of the hydraulic device casing and has expelled the fluid from this volume, the actuator piston, performing its reverse, engine-valve opening movement, the vacuum created above the rod exiting the space above it will counteract the opening of the engine valve. This will also delay the opening of the engine valve, creating a secondary deceleration effect and showing that the braking device described herein operates as a reciprocating unit, decreasing the speed of the gas-exchange valve as it closes, shortly before the valve contacts its seat, and, secondly, as the valve opens. The secondary deceleration of the gas exchange valve is undesirable.
The changing decelerating effect of the braking device and the secondary deceleration of the gas-exchange valve when opening and closing are undesirable and result in reduced valve actuator operation security and, ultimately, deteriorated engine operation.
The use of solenoid valves to control the fluid in the pneumatic portion of the actuator will increase its operational noise levels. Increased noise is also caused by the fluid (gas) exiting the cylinder volume of the actuator via the envisaged complex-shaped and relatively long discharging channel.
The need for additional devices to drive the pneumatic part (a compressor or a pressurized gas-tank) for the hydraulic and the pneumatic parts used in the actuator complicates the operation and size of this mechanism.